Combination single and dual stage compressor



Nov. 15, 1966 SAHLE ET AL 3,285,500

COMBINATION SINGLE AND DUAL STAGE COMPRESSOR Filed May 25, 1964 5 Sheets-Sheet 1 TONNAGE (BTU/h) I COOLING CAPACTTY E A l8) B RATTO 4 PRESSURE DUAL RATIO /\.44 OUTPUT STAGE \NPUT CONTROL COMPRESSOR l J, J

INVENTORS KNUTE SAHLE ALWTN b. NEWTON & ATTO EY Nov. 15, 1966 K, AHLE ET AL 3,285,500

COMBINATION SINGLE AND DUAL STAGE COMPRESSOR Filed May 25, 1964 -5 Sheets-Sheet 2 p' 1 Q 1 Q FIGS T CYLS I? OUTPUT II V 40 FROM CYLINDERS INVENTORS KNu'rE SAHLE 45 Am: B. NEWTON FROM B CYLINDERS A ORNEY OUTPIUT COMBINATION SINGLE AND DUAL STAGE COMPRESSOR Filed May 25, 1964 K. SAHLE ET AL Nov. 15, 1966 5 Sheets-Sheet 3 FIG? 6 N U o o N mam Wm w W A 55 E WW NL KA United States Patent 3,285,500 COMBINATION SINGLE AND DUAL STAGE COMPRESSOR Knute Sahle and Alwin B. Newton, York, Pa., assignors to Borg-Warner Corporation, Chicago, 111., a corporation of Illinois Filed May 25, 1964, Ser. No. 369,731 Claims. (Cl. 230-40) This invention relates to a compressor and more particularly to a dual stage compressor including a sensing means to change the mode of operation from single stage compression to that of dual stage compression upon the occurrence of a certain pressure ratio as sensed by an appropriate sensing means, the sensing means also forming a part of this invention.

Compressors of many types are used for performing a great variety of tasks. In the refrigeration art, as exemplary, compressors are used to compress a suitable working substance during each cycle, the compressor transferring to the working substance energy in the form of mechanical movement after the working substance has ex panded to produce its cooling eflect. In many applications, it is desirable that compression be effected in two stages, each stage being in series with an adjacent stage whereby the output of one stage becomes the input for the adjacent stage. While advantageous in certain instances, in other instances such a mode of operation of a plurality of stages suffers the disadvantage of a diminution in capacity of the units, in that a lesser number of cubic feet per minute may be compressed. Such a mode would be in distinction to another mode of operation wherein a plurality of compressional stages may be placed in parallel between an input and an out-put. A parallel mode of operation, while enjoying the advantage of initial greater capacity, suffers the disadvantage of limited compression since only a single stage is effectively employed to compress the working substance.

In certain applications, the most economical mode of operation of a plurality of stages dictates that the stages be operated in parallel amongst one set of conditions and be operated in series at another set of conditions.

Accordingly, it is an object of the present invention to provide a compressor having a plurality of stages wherein the determination of the mode of operation of a number of stages is the function of the ratio of the lowest or input pressure to the highest or output pressure.

It is a further object of the present invention to provide a compressor having a plurality of compressional stages to vary the mode of connecting the stages so as to yield the most economical mode of operation as a function of a ratio between two certain pressures.

It is a further object of the present invention to provide a pressure ratio sensing and controlling means to sense the ratio between two pressures and to thereby effect a control of a mechanism associated therewith.

These and other objects of the present invention will become apparent from the following.

In the drawings:

FIGURE 1 illustrates the behavior of two compressors when acting in parallel and when acting in series, the curves showing the relation between cooling capacity vs. pressure ratio for a refrigeration system.

FIGURE 2 is a partially schematic view of a plenum chamber of a (six-cylinder) compressor.

FIGURE 3 is a partially schematic vieW of certain elements of FIGURE 2.

FIGURE 4 is a schematic view showing the cooperation between the compressional stages and a pressure ratio controller.

FIGURE 5 is a partially schematic cross-sectional view of the pressure-ratio senser and controller of the present invention.

Patented Nov. 15, 1966 FIGURE 6 is a partial cross-sectional view of the compressor of this invention showing the crankcase.

FIGURE 7 is a partial cross-sectional view of one element shown in FIGURE 6.

FIGURE 8 is a partial cross-sectional view of another element shown in FIGURE 6.

Referring now to FIGURE 1 of the drawings, the illustrated curves represent the behavior of two fluid compression means when coupled to act in series and when coupled to act in parallel. For convenience, with reference to this figure, each compression means may be thought of as comprising a cylinder and reciprocating piston therein with associated input and output valves. Also, the independent fluid compression means will frequently be referred to herein as the first and second stages although it is understood that, technically, they function as two stages only when coupled in series. When coupled in parallel they operate as parallel, single stage compression means which have a common inlet and a common outlet. The cooling capacity of the stages (the invention exhibiting particular utility in the refrigeration art) is plotted as the dependent variable against the ratio of output to input pressure as the independent variable. The curve A-B represents the capacity variation with pressure ratio when the two compressional stages are coupled in series. The curve CD represents the variation in ca pacity when they are connected in parallel. It will be seen that the curves cross at a point E corresponding to a particular pressure ratio.

From an inspection of FIGURE 1, it will be observed that most efficient operation of the available compressional stages, i.e., that mode of operation yielding the greatest cooling capacity, would be obtained when they are operated in parallel from relatively low pressure ratios up to the point E and thence operated in series from a pressure ratio corresponding to the point E onwards. Thus the optimum operating path would be CE-B.

In a manner now to be set forth, the practice of the present invention allows for operation along the optimum path by changing from one mode of operation to another mode of operation at a predetermined pressure ratio, i.e., the pressure ratio corresponding to the point E of FIG- URE 1.

Referring now to FIGURE 2 of the drawings, the numeral 10 denotes generally the outline of certain plenum chambers associated with the suction side of the cylinders in a compressor according to this invention. The numeral 11 denotes portions of a wall or septum generally upstanding from the top portions of the block or body of a compressor. It will be understood that in practice the wall portions 11 need not be coplanar and hence FIGURE 2 may be regarded as a projection on a two-dimensional surface or as an actual unfolding, as in the manner of unfolding a cardboard box so that it lies on a flat surface. They are here shown as coplanar for clarity. The numeral 12 denotes an aperture in the wall 11 and through which is adapted to pass the (working) substance to be compressed by the device. The numeral 13 indicates another wall portion and is provided with an aperture 14 in fluid communication with aperture 12. The numerals 15 each denote a cylinder, there being four such in number, by way of example only, with the input or low pressure sides thereof facing the viewer. The number 16 denotes another wall portion, preferably integral at its ends with wall 11, and surrounds the input ends of two additional cylinders 17. The numeral 18 denotes a continuous flow path as in the form of a conduit from the output of the cylinders 15 to a fitting 37 (to be described later), the fitting 37 forming a portion of an ensemble of elements shown in FIGURE 3.

Referring now to FIGURE 3 of the drawings, combination valve means and valve actuating means V is provided for selectively placing the groups of cylinders 15 and 17 in series or in parallel. The numeral 20 denotes an insert, as in the manner of a plug, in an aperture in wall 16. A seal 21, formed in a peripheral groove, surrounds the insert to maintain a fluid impervious joint. The numeral 22 denotes a central aperture extending through the plug, the former being provided with an upstanding seat 23 around one edge. closed at one end and whose other end includes an integral flange. The cylinder is spaced from plug 20 by a plurality of angularly spaced hollow sleeves 25, the plug 20 in turn being spaced from fitting 37 by a plurality of angularly spaced sleeve members 26, with a bolt 27 passing through the sleeves 25 and 26. It will be understood that while only one such bolt 27 is illustrated there are a plurality through the periphery of the flange of cylinder 24. The numeral 28 denotes a piston adapted to recipro- The numeral 24 denotes a cylindercate within cylinder 24 and is borne against on its exterior end by one end of a spring 29. The numeral 30 denotes the volume between the closed end of cylinder 24 and the interior face of piston 28. A bolt 31 extends completely through the piston 28 and is surrounded for a major portion of its length by a sleeve 32, the sleeve and bolt passing through an aperture in a plate 33 which, eX-

cept for the central aperture therein, completely closes the open mouth of cylinder 24. The other end of spring 29 abuts plate 33. The numeral 34 denotes a movable valve element, sometimes referred to hereinafter as a plug, mounted for reciprocation with piston 28 by virtue of its shown connection with bolt 31 and is faced on both sides with a suitable sealing material such as disc elements 35 of Teflon. A pair of washer elements 36 cooperate with the bolt 31 and sleeve 32 to hold the plug assemblage together.

The numeral 37 denotes a pipe fitting in the general form of a T having the opening of branch 38 provided with a rounded, integral seat 39. It will be observed that in the illustrated position of the piston 28, the right Teflon disc 35 rests upon and forms a fluid tight seal with seat 23 of aperture 22 and that when, in the leftmost position (not illustrated) the left Teflon disc 35 will abut and form a fluid-tight seal with seat 39. It will be understood that when the plug 34 is not seated around aperture 22, there exists fluid communication between the zone adjacent the left portion of plug member 20 and the space into which cylinder 24 projects.

Referring now to FIGURES 2, 3 and 4, the output of cylinders 15, passing through conduit 18, is fed to one end of T-fitting 37, the other head end of which leads directly to one side of a combination check and stop valve 40, the other side of the latter being secured to one head end of another T-fittiu-g, 41 whose leg 42 is secured to elbow 43, into which the output of cylinders 17 is fed. The free head end of T-fitting 41 (the upper end of FIG. 3) represents the discharge or output from the compressor. When operating as a dual stage compressor, cylinders 15 may be regarded as the first stage of compression and cylinders 17 as the second stage.

Referring now to FIGURE 4 of the drawings, the operation of that portion of the compressor according to the present invention which has been above described will now be given, before commencing the description of the operation of the pressure ratio controller of FIGURE 5 which also forms a part of the present invention.

The numeral 44 denotes a pressure ratio controller having both a low and a high pressure input. These inputs are derived from the compressor and in practice are carried out by thin tubes (denoted by LP. and H.P.) secured, respectively, to the low and high output portions of the dual stage compressor. Depending upon the ratio of the high pressure (second stage discharge pressure) to the low pressure (first stage input pressure) fed to the controller 44, a solenoid valve 45 will either be energized so as to supply conduit 46 (leading into volume 30 within cylinder 24) either with low pressure fluid or with high pressure fluid,

both pressures being taken from the dual stage compressor, as by thin tubes.

Assuming all elements above described to be at rest, initiation of the operation of the dual stage compressor finds the piston 28 in the illustrated position due to the action of spring 29. As the action of thecornpressor commences, the ratio of the high pressure output of the compressor to the suction (first stage input) will, at least temporarily, be less than a predetermined amount. Pressure ratio control 44, sensing this ratio, actuates solenoid valve 45 so that the output of cylinders 17, the output of the second compressional stage, is fed through conduit 46 into space 30 of cylinder 24. This causes plug 34 to move to the left against the action of spring 29 until it abuts seat 39. The flow of the compressor input (here see the solid curved arrows of FIGURE 2) is through apertures 12 and 14 and to the input sides of the four cylinders 15. The plug 34 having seated against seat 39, another portion of the input of the compressor follows the dashed curved arrows of FIGURE 2 and thus enters the side of wall or septum portion 16 which surrounds the input sides of the two cylinders 17. The output of the four cylinders 15 is fed through conduit 18 to one end of T-fitting 37 where it communicates with the aperture in branch 38 and exerts a force against the left (FIGURE 3) face of plug 34 urging it to the right, this force being resisted however by a larger force urging it to the left by the same output through conduit 46 acting on the interior face of piston 28. It will be observed that the area of the interior face of piston 28 is greater than the area of the apertures which plug 34 close and hence the piston will remain seated on seat 39. The output from the four cylinders 15, through conduit 18, thus passes completely through T-fitting 37, none escaping through the branch 38 thereof, and into one side of combination check and stop valve 40. The construction permits the flow of gas to pass one way only, towards T-fitting 41. The output of the two cylinders 17 is fed into elbow 43, hence into T-fitting 41, and thus joins the discharge of the fourcylinders 15 in T-fitting 41 in parallel, the total output discharging through the free head end of fitting 41. It will be noted that after the output of cylinders 17 initially expands into volume 30, further expansion does not take place and thereafter the entire output of cylinders 17 is to elbow 43. This represents the parallel or first mode of operation of the stages, and occurs at those ratios of output to input pressure to the "left of point B of FIGURE 1.

If, during the operation of the compressor, the output pressure should increase or, as would -be more likely in the case of a refrigeration system, the suction or input pressure should decrease, the ratio of output pressure to input pressure would increase to and beyond the point B of FIGURE 1. At and above a pressure ratio cor responding to E of FIGURE 1, pressure ratio control 44 causes solenoid valve 45 to feed the suction or input pressure of the first stage through conduit 46 to space 30 within the cylinder 24. The output of cylinders 15, fed through conduit 18, now acts against the left face of plug 34 and moves the plug and piston 28 to the right, back to the illustrated position, because the product of the (first stage) output pressuretimes the area of the left face of plug 34 is now greater than the product of the suction pressure (first stage input) in area 30 and the interior face of piston 28. It will be understood that this movement of plug 34 from seat 39 back to seat 23 is quite rapid, and, assisted by spring 29, there is no opportunity for pressure coming from branch 38 to also act on the right face of plug 34 to maintain the latter in an intermediate position between seats 39 and 23. Thus, the plug is always, for practical purposes, positioned on one or the other of these seats. In order to preclude hunting, the pressure ratio control 44 or the valve 45 may be provided with a time-delay means, so that a pressure ratio which dictates a change would have to be continuous over a predetermined length of time.

With the plug 34 on the right seat 23 as illustrated, the incoming gas (see FIGURE 2) is precluded from following the path of the curved, dashed arrows and the output of cylinders 15 through conduit 18 passes through branch aperture 38, past the left face of plug 34 and into the input sides of cylinders 17. The output of cylinders 17, led to elbow 43, passes into T-fitting 41 through branch 42 and bears against the other side of combination check and stop valve 40. Because of the extra stage of compression, the combination check and stop valve 40 precludes any of the first stage output from passing therethrough, the series sum of the stages bearing against the top side thereof (FIGURE 3) and the first stage only acting against the bottom side. The entire output of the low stage cylinders 15 of the compressor thus passes into the input sides of cylinders 17 and thence out through elbow 43 to the output of the compressor. This represents the series or second mode of operation of the stages and occurs at those ratios of output to input pressure to the right of point E of FIGURE 1. It will be observed that the first and second stage output pressures, as well as their input pressures, are the same in the parallel mode of operation.

From the above, it will be seen that the ideal path of operation, CE-B of FIGURE 1, is attained.

The details of the solenoid valve 45 of FIGURE 4 have not been illustrated, since any one of a great number of types of such valves may be employed. It is only necessary to recall that it is a three-way valve, preferably actuated by a solenoid, with the common output thereof coupled to conduit 46, the valve connecting either the input of the first stage suction or the output of the second stage discharge to conduit 46. In practice, the solenoid valve communicates with the input of the first stage suction and the output of the second stage discharge by small tubular conduits.

Referring now to FIGURE of the drawings, a description of the pressure ratio controller of this invention Will now be set forth. The numeral 47 denotes generally the pressure ratio controller and comprises a closed housing 48 which may be formed of sheet metal or the like. The numerals 49 and 50 denote integral tubular extensions of bottom portions of the housing, the extensions being coupled at their lower ends to conduits .51 and 52. Conduit 51 is in fluid communication with the suction or input side of the first stage of the compressor and conduit 52 is in fluid communication with the output of the second stage of the compressor.

The numerals 53 and 54 denote generally cylindrical bellows whose upper ends are open and flared and suitably affixed to the periphery of extensions 49 and 50. The lower ends of the bellows are closed as by discs denoted by numerals 55 and 56, the latter closures carrying, respectively, push rods 57 and 58. The upper ends of the push rods conically terminate and bear against inserts 57a and 58a to thus define needle bearings. The numerals 59 and 60 denote two bellcrank levers pivoted, respectively at 61 and 62 to the frame 48. Stop or abutment elements 59a and 60a are carried by frame 48 to preclude inward movement of their associated levers more than part-way toward the mid point 80 of spring 79. The numerals 63 and 64 denote extensions of cranks 59 and 60, respectively, and are placed near brackets 65 and 66, the latter suitably atfixed as by welding to the frame 48. Each bracket carries a threaded fastener or bolt 67 and 68, respectively, whose end carries a nut threaded thereto, the nut of each suitably aflixed to springs 69 and 70. Th numeral 65a denotes one of a pair of knife-edge supports, diametrically spaced on opposite sides of the aperture on bracket 65 through which bolt 67 passes. The numeral 65b denotes a washer one side of which is provided with mating recesses therein to receive the knife edges. The head of bolt 67 abuts the other side of 6 washer 65b. Thus the bolt 67 is free to rotate about its own :axis relative to washer 65b and is also free to pivot about the knife edges 65a about an axis perpendicular to the viewer upon rotation of lever 59. Bracket 66 is provided with a similar structural arrangement. The ends of the springs 69 and 70 are adapted to pass through and onto extensions 63 and 64. It will be observed that th longitudinal axes of bolts 67 and 68 pass directly through pivots 61 and 62 in the illustrated neutral position. The numerals 71 and 72 denote other bolts, each carrying, respectively, a nut 73 and 74 affixed to one end of springs 75 and 76. The other ends of the latter are affixed to the upper ends of the bell crank arms 59 and 60.

The numeral 77 denotes a lever pivoted at 78 to the frame 48 and which carries a spring 79 whose midportion 80 passes through an aperture in the lever 77. Alternatively, a pin or other suitable fastening element may fix the midpoint of spring 79 to lever 77. Further, spring 79 may be in two portions, each lying on one side of lever 77 The ends of spring 79 bear against facing sides of bell cranks 59 and =60, :as by the provision of small notches therein into which the ends are positioned. The numeral 81 denotes an electrical contact secured to the end of lever 77, the contact carrying a thin flexible conductor 82 secured to binding post 83, the latter being suitably insulated from frame 48. The numeral 84 denotes another electrical contact, here carried by a bracket, the latter being in contact with another binding post 85 also suitably insulated from frame 48.

In operation, the input pressure of the first stage of the compressor bears against the closure 55 of bellows 53 thus urging the right hand bell crank 59 in a counterclockwise direction about its pivot point 61. The output pressure of the secondstage of the compressor (output pressure of cylinders 17) is connected to conduit 52 and acts against closure 56 of bellows 54, urging left hand bell crank 60 in a clockwise direction about its pivot 62. Th pressure introduced into conduit 52 being, usually, substantially higher than the pressure in conduit 51, the size of the bellows 54 is generally smaller than the size of bellows 53 and accordingly the cylindrical extensions 49 and 50 are of different sizes. A spring S may be employed in conjunction with bellows 53 yielding additional force urging bell crank 59 counter-clockwise with sufiicient force to provide a net counter clockwise component even under conditions wherein the pressure in conduit 51 is below atmospheric pressure. g

It will be observed that the resultant motion of lever 77 will be the summation of the oppositely acting motions of the bell cranks 59 and 60, these motions being transferred to the lever by means of the spring 79. It will further be observed that the making or the breaking of electrical contacts 81 and 84 will depend upon the movement of lever 77. Electrically actuated solenoid valve 45 (see FIGURE 4) carries two terminals which are connected to contact posts 83 and 85 and it will be thus observed that the actuation of the solenoid valve will depend upon the ratios of the pressures led to the control 47 through conduits 51 and 52. Springs 75 and 76 may be varied so as to exert dilferent tensions on the bell crank arms 59 and 60 by adjustment of bolts 71 and 72.

It will be further observed that each of the deflecting movable elements associated with levers 59 and 60 including springs 76, bellows 54, and spring 79 with respect to lever 60 and spring 75, spring S, bellows 53, and spring 79 with respect to lever 59, all impose forces on the levers 59 and 60 which vary in proportion to the movement of the levers 5'9 and 60 from any previously occupied position. In other words, the spring and bellows system of each lever 59 and 60 has its own definite spring rate and, therefore, will move a predictable amount for each increment of additional force imposed against the surfaces 55 and 56 of bellows 53 and 54. The functions of springs 69 and 70 is to impose additional forces which are proportional (for the small angles here encountered) to deflection or movement of levers 59 and 60, but which do not affect the mean setting of springs 75 land 76. In the position shown for levers 59 and 60, any force exerted by springs '69 and 70 merely urges the said levers with greater force against their respective pivots 61 and 62 without affecting their motion, since a straight line relationship exists between the pivots, the points of spring attachment on their respective levers, and the points of anchorage of the springs at 65 and '66. Any motion, however, of either lever disturbs this in-line relationship and allows the corresponding spring 69 or 70 to exert a restoring force proportional to the departure from the in-line position which force tends to return the lever to the in-line position. Since this restoring force is proportional to the tension initially imposed on springs 69 and 70 it can be varied at will by adjustment of bolts 67 and 68. To adjust the pressure ratio control 47, spring 75 is adjusted to the expected midpoint of the range of suction pressures to be encountered by the compressor for any specific application on which it is applied. Spring 69 is then adjusted by means of screw 67, so that any given desired motion, for example A", at the upper end of lever 59 will be obtained when imposing first the maximum and then the minimum anticipated suction pressure. Spring 76 is likewise adjusted to the mid-point of the range of discharge pressures anticipated for the particular application of the compressor, and spring 70 is adjusted by means of screw 68 to secure a predetermined motion of the upper end of arm 60, when first the lowest and then the highest discharge pressure is applied to conduit 52. When so adjusted the upper end of each lever 59 and 60 tends to reach a definite position for each momentarily existent pressure in conduits 51 or 52.

The motion of point 80 on spring 79 is affected by movement of either lever 59 or 60 and serves to measure the ratio between pressures in conduits 51 and 52. Thus contacts 81 and 84 are closed or open only in response to the relative motions of levers 59 and 60 and not to their absolute positions. By choosing the adjustments of springs 69 and 76 so that the motions of the levers 59 and 60 are proportional to the squares of the desired ratios of the range of one pressure compared to the other, contacts 81 will respond to compression ratio between the high and low side of the compressor and not to absolute pressure.

For example, if the suction pressure is expected to range from 7# to 95# absolute, a range of 88#, and the discharge pressure is expected to range from 165# to 415# absolute, a range of 250'#, the square of the ranges becomes (250/ 88)' =8. Then springs 69 and 70 are adjusted so that when a given pressure change moves -lever 59 for any given distance such as .001" it will require 8 times this pressure change to move lever 60 a like distance or .001". 1

'Turning now to FIGURES 6, 7 and 8, the numeral 100 denotes generally the crankcase of the compressor of this invention and the numeral 101 designates the first compressional stage input of the compressor, sepanated from the interior of the crankcase 100 by wall portions 102. The numeral 103 denotes a bath of lubricating oil for lubricating the various portions of the machine and numeral 104 denotes a plurality of turns of thin and preferably flexible tubing defining a coil opened at one end and the other end is connected to a fitting 105. The fitting 105 is adapted to be coupled to a source of liquid refrigerant and the illustrated gas bubbles emanating from the free end of the coil 104 upwardly through the lubricant 103 denotes the passing of the refrigerant from the liquid to the gaseous state. Heat from the relatively warm oil is transferred to the coil, causing the change of state. Thus, liquid refrigerant fed through fitting 105 is in heat exchange relationship with lubricant 103 and serves to preclude an extremely high temperature of the lubricant 103.

(3 The numeral 106 denotes a fitting preferably threaded into wall 101 as indicated and FIGURE 8 shows this fitting in detail. The numeral 107 denotes a conduit leading from fitting 106 to an oil pump. The numeral 108 denotes a coupling element having one portion exteriorly threaded and mounted through an aperture in wall portion 102. The numeral 109 denotes a movable piston adapted to reciprocate within a bore in 108. The numeral 110 denotes one of a plurality of apertures in the wall of connector 108 and the numeral 111 denotes an axially extending aperture in the connector 108. The numeral 112 denotes a compressor coil spring which normally urges the piston 109 in the illustrated position. The numeral 113 denotes another coupling element threaded into coupler 108. In the illustrated position of FIGURE 8, the position of piston 109 permits fluid communication between the crankcase chamber and the volume 101 of the low stage suction of the compressor. This enables any difference in pressure between the interior of the crankcase and the low stage suction to be zero just before operation of the compressor is begun. This initial lack of pressure differential enables easier starting of the compressor since there is generally higher pressure within the crankcase than within chamber 101 and this higher pressure, if not relieved at compressor shutdown, would result in a pressure differential which would make the initial starting torque requirements for the compressor prime mover greater than with no pressure differential. After the operation of the compressor is commenced, the fluid pressure through line 107 (see FIGURE 6) acts on the right hand face of piston 109 urging the conical extension thereof up into aperture 111, thus closing it, and thus sealing off the fluid communication between these chambers.

In addition to the function above described, the valve 106 may be placed in wall 102 so that it is at the lowermost portion of the suction chamber and would then occupy a position at the lowest point of the Y-shaped configuration at the left portion of FIGURE 6 adjacent element 106. In such a position, any oil in the low stage suction area at shutdown would gravitate to the lowest portion in the suction chamber 101 and pass through apertures 111 and 110 into the crankcase. This would preclude the collection of a large quantity of oil which might be saturated with refrigerant from collecting in the low stage suction chamber where it could slug the compressor on start-up with possible resulting valve damage.

Referring now to FIGURE 7, the numeral 115 denotes a valve positioned in wall 102 and includes a generally cylindrical portion 116 having one end thereof closed by a screw threaded plug 117. One end of the plug 117 bears against compression spring 118, the latter having one end which bears against spherical element 119. The ball 119 is normally urged against seat 120, the seat also defining an aperture which communicates, when open, with the interior of the crankcase 100. Passageway 121 communicates with the interior of the cylinder 116 and, when the pressure within the crankcase is greater than a predetermined amount, corresponding to the force of spring 118, the ball 119 is urged against the force of the spring and establishes fluid communication between the interior of the crankcase and the low stage suction chamber 101. By this arrangement, a predetermined pressure differential during compressor operation may be maintained between the interior of the crankcase and plenum chamber 101. Thus, blow-by gas from the high stage pistons, together .with the vaporized refrigerant from coil 104, are precluded from causing the pressure within the crankcase to assume extremely high values. Such vapors, at proper pressure values, are bled through the valve 115 to the input chamber 101. The valve 115 further controls the loading.

on the piston pins between the pistons and their connecting rods such as piston 17a and piston rod 17b of FIG- URE 6. It is apparent that by controlling the pressure within the crankcase 100, the force acting on the bottom of a piston such as 17a can be controlled. In many applications, load reversal is necessary for optimum lubrication results and the pressurizing and pressure controlling of such pressure by valve 115 is realized here. By suitably loaded spring 118 of valve 115 by adjustment of plug 117, a pressure condition during compressor operation within crankcase 100 may be realized which will insure a load reversal on the piston pins.

We claim:

1. A compressor comprising first and second fluid compression means each having suction and discharge sides; means defining first and second suction plenums associated with the suction sides respectively of both said compression means, said plenum defining means including a septum dividing the respective plenums, said septum having a first aperture therein to provide fluid communication between said first and second plenums, said plenum defining means further including a wall section of said second suction plenum having a second aperture formed therein; first conduit means interconnecting the discharge side of said first fluid compression means to a fluid outlet; second conduit means interconnecting the discharge side of said second compression means with said first conduit means at a juncture where the discharge from both compression means are combined; third conduit means defining a fluid passage interconnecting said second aperture with said first conduit means upstream from the juncture of said first and second conduits; a one-way check valve upstream from said juncture and downstream from said third conduit means; valve means movable from a first position wherein said first aperture is open and said second aperture is closed, thereby permitting communication between said first and second suction plenums, to a second position wherein said first aperture is closed and said second aperture is open, thereby permitting the discharge from said first compression means to flow into said second plenum; and valve actuating means operatively associated with said valve means responsive to the pressure differential between the suction side of said first compression means and the discharge side of said second compression means.

2. A compressor as defined in claim 1 wherein said valve actuating means includes a piston reciprocatively movable within a cylinder; a three-way valve actuated in response to said pressure differential, said three-way valve adapted to selectively apply the suction side pressure of said first compression means to said cylinder, or alternatively, the discharge side pressure of said second compression means to said cylinder.

3. A compressor as defined in claim 2 including means for sensing the pressure diflerential between the suction side of said first compression means and the discharge side of said second compression means; and a solenoid actuating said three-way valve, said solenoid being operated by said pressure differential sensing means.

4. A compressor comprising first and second fluid compression means each having suction and discharge sides; means defining first and second suction plenums associated with the suction sides respectively of both said compression means, said plenum defining means including a septum dividing the respective plenums, said septum having a first aperture therein to provide fluid communication between said first and second plenums, said plenum defining means further including a wall section of said second suction plenum having a second aperture formed therein; first conduit means interconnecting the discharge side of said first fluid compression means to a fluid outlet; second conduit means interconnecting the discharge side of said second compression means with said first conduit means at a juncture where the discharge from both compression means are combined; third conduit means defining a fluid passage interconnecting said second aperture with said first conduit means upstream from the juncture of said first and second conduits; a one-way check valve upstream from said juncture and downstream from said third conduit means; valve means including a valve element reciprocatively movable between a first seating position over said first aperture to block communication between said first plenum and said second plenum and a second seating position over said second aperture to provide fluid communication between said first and second plenums; means for actuating said valve element, said means including a piston reciprocatively movable within a cylinder, the area of said piston being greater than the area of said second aperture; means for selectively applying suction pressure of said first compression means or the discharge pressure of said second compression means to a fluid working space behind said piston, whereby, when the discharge pressure of said second compression means is applied to said working space the valve will close said second aperture to effect parallel operation of said first and second compression means, and when the suction pressure of said first compression means is applied to said working space the discharge pressure of said second compression means acting against the valve element will cause it to close said first aperture and effect .series operation.

5. A compressor as defined in claim 4 including a pressure ratio control means to sense the ratio of the discharge pressure of said second compression means to the suction pressure of said first compression means; a three-way valve, said three-way valve being operated by said pressure ratiocontrol means, said three-way valve selectively applying the suction pressure of said first compression means to said working space when the pressure ratio is greater than a predetermined ratio, and applying the discharge pressure of said second compression means to said working space at a pressure ratio less than said predetermined ratio.

References Cited by the Examiner UNITED STATES PATENTS 2,246,932 6/1941 Collins 23040 2,329,401 9/ 1943 Le Valley 23029 2,391,486 12/1945 Smith 23040 2,656,970 10/1953 Pfiefer 230-29 2,761,387 9/1956 Gaubatz 10310 2,971,070 2/ 1961 Snoberger 20083 2,983,432 5/1961 Tupper 23040 3,194,915 7/1965 Anderson 200453 3,211,365 10/1965 Phelps 103-207 MARTIN P. SCHWADRON, Primary Examiner.

SAMUEL LEVINE, MARK NEWMAN, Examiners.

W. L. FREEH, Assistant Examiner. 

1. A COMPRESSOR COMPRISING FIRST AND SECOND FLUID COMPRESSION MEANS EACH HAVING SUCTION AND DISCHARGE SIDES; MEANS DEFINING FIRST AND SECOND SUCTION PLENUMS ASSOCIATED WITH THE SUCTION SIDES RESPECTIVELY OF BOTH SAID COMPRESSION MEANS, SAID PLENUM DEFINING MEANS INCLUDING A SEPTUM DIVIDING THE RESPECTIVE PLENUMS, SAID SEPTUM HAVING A FIRST APERTURE THEREIN TO PROVIDE FLUID COMMUNICATION BETWEEN SAID FIRST AND SECOND PLENUMS, SAID PLENUM DEFINING MEANS FURTHER INCLUDING A WALL SECTION OF SAID SECOND SUCTION PLENUM HAVING A SECOND APERTURE FORMED THEREIN; FIRST CONDUIT MEANS INTERCONNECTING THE DISCHARGE SIDE OF SAID FIRST FLUID COMPRESSION MEANS TO A FLUID OUTLET; SECOND CONDUIT MEANS INTERCONNECTING THE DISCHARGE SIDE OF SAID SECOND COMPRESSION MEANS WITH SAID FIRST CONDUIT MEANS AT A JUNCTURE WHERE THE DISCHARGE FROM BOTH COMPRESSION MEANS ARE COMBINED; THIRD CONDUIT MEANS DEFINING A FLUID PASSAGE INTERCONNECTING SAID SECOND APERTURE WITH SAID FIRST CONDUIT MEANS UPSTREAM FROM THE JUNCTURE OF SAID FIRST AND SECOND CONDUITS; A ONE-WAY CHECK VALVE UPSTREAM FROM SAID JUNCTURE AND DOWNSTREAM FROM SAID THIRD CONDUIT MEANS; VALVE MEANS MOVABLE FROM A FIRST POSITION WHEREIN SAID FIRST APERTURE IS OPEN AND SAID SECOND APERTURE IS CLOSED, THEREBY PERMITTING COMMUNICATION BETWEEN SAID FIRST AND SECOND SUCTION PLENUMS, TO A SECOND POSITION WHEREIN SAID FIRST APERTURE IS CLOSED AND SAID SECOND APERTURE IS OPEN, THEREBY PERMITTING THE DISCHARGE FROM SAID FIRST COMPRESSION MEANS TO FLOW INTO SAID SECOND PLENUM; AND VALVE ACTUATING MEANS OPERATIVELY ASSOCIATED WITH SAID VALVE MEANS RESPONSIVE TO THE PRESSURE DIFFERENTIAL BETWEEN THE SUCTION SIDE OF SAID FIRST COMPRESSION MEANS. 